Elastic fluid compressor



y 24, 1955 J. M. STEPHENSON 2,709,036

ELASTIC FLUID COMPRESSOR 3 Sheets-Sheet 1 Fig. I.

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ELASTIC FLUID COMPRESSOR Filed June 24, 1949 5 Sheets-Sheet 3 FIG. ll.

United States Patent ELASTIC FLUID COMPRESSOR John M. Stephenson, Manchester, England, assignor to Power Jets (Research and Development) Limited, London, England, a British company Application June 24, 1949, Serial No. 101,220

Claims priority, application Great Britain July 6, 1948 3 Claims. (Cl. 230-209) This invention relates to elastic fluid compressors of the dynamic as distinct from the positive displacement type, which includes, for example, rotary axial flow and centrifugal compressors, and is concerned with those designed for multi-stage operation and having cooling provided between the stages.

It is well known that dynamic compressors, when operating under certain conditions of mass flow of fluid and pressure ratio, are subject to the occurrence of a phenomenon known as surging in which the fluid flow, normally uni-directional, becomes unstable and oscillates.

The limiting condition between stable and unstable operation may be defined, in a graphical representation of the characteristics of a dynamic compressor, by a curve known as the surge line; a typical example of such a characteristic is shown in Figure 1 of the accompanying drawings, which represents diagrammatically the pressure ratio R (ordinates) plotted against the corresponding mass flow M (abscissae) of fluid through the compressor. The full line S is the surge line, all conditions of flow lying to the right thereof being stable and all to the left unstable. It is desirable that the compressor, when constituting a component of a heat engine or other plant, should operate closely to the surge line but still have an inherently stable flow; such operation would be represented by the broken line 0 in Figure 1, which remains in the area of stable operation over a wide range of fluid mass flow or compressor load.

It is, however, impracticable to arrange, simply by suitable design of the compressive elements, for the compressor to have such a desirable operating line; in fact only one point on the line could be aimed at in designing the unit; the ultimate operating line would be dictated by the conditions in the plant as a whole and by the ability of the compressor to adjust itself to variations in those conditions. This latter feature is affected by the actual structural form of the compressor; e. g. a multi-stage rotary compressor having its stages all on one shaft is less flexible in this respect than a similar compressor having its stages mounted on independent shafts, although the latter is, of course, more complex structurally. Thus certain dynamic compressors, whilst having otherwise desirable features, may have, when constituting a component of a plant, a relatively poor operating line; a typical example of such a case is illustrated by the broken curve 01 in Figure 1. As will be seen, this line intersects the surge line obliquely and the compressor thus operates with stability only over a comparatively limited range of mass flow or load. Such an operating line is typical, for instance, of a multi-stage rotary compressor having its stages all in one shaft, as commonly used in gas turbine power plants, and constitutes a major limitation of the type.

Furthermore, when a compressor of the kind indicated is provided with means for cooling the fluid at intermediate stages of compression, as is generally desirable in order to achieve optimum efliciency, its characteristic tends to be further adversely affected and an operating 2,7050% Patented May 24, 1955 line of the type illustrated by the broken line 02 in Figure 1 might be obtained; this line intersects the surge line more obliquely than does 01 and the compressor load range for stable operation is further restricted.

It should be understood that the point of intersection A of the respective operating lines 0, 01, Oz, in Figure 1 would correspond to a full load design condition common to each of the compressors, and also that, in the case of the compressor with inter-stage cooling, the cooler is normally so designed as to effect a temperature drop in the fluid delivered at the full load design condition suflicient to return it to its temperature at inlet, so simulating isothermal compression.

The cooler, however, in addition to decreasing the fluid temperature, increases the density of the fluid by a corresponding amount, i. e. the density varies in inverse relation to the temperature, and the subsequent stage or stages of the compressor would be designed with a capacity corresponding to this changed fluid density.

When, however, a dynamic compressor is run at part loads, in addition to the reduction in fluid mass flow, the pressure ratio, and consequently the temperature ratio, of compressed to uncompressed fluid are reduced from their full values; the inter-stage cooler, therefore, is required to provide only a reduced temperature change in order to restore the fluid to its original temperature, and hence, the ratio of densities of the cooled to the uncooled fluid is reduced as compared with the ratio at full'load.

The efiect of this reduction in density ratio with load reduction is to cause the velocity ratio of fluid flow through the compressor to depart from the design condition; hence the limited range and proneness to surging of such a compressor. The invention has for its basis the realisation that if a reduction in load were to be accompanied by an increase in the density ratio of the cooled to the uncooled fluid, the limitation of range normally associated with inter-cooled compressors would be avoided, and even the load range of an uncooled compressor may be improved upon (as illustrated, for example, by the curve 03 in Figure 1).

The invention therefore proposes, in a plant incorporating a multi-stage elastic fluid compressor having provision for cooling the fluid between successive stages, the method of varying the degree of cooling applied with varying compressor loads such that the ratio of the densities of the fluid after and before cooling respectively is increased when the compressor load decreases and is decreased when the compressor load increases.

In general the desired effect is that the inter-stage cooling shall be progressively less effective as the compressor load is increased.

Therefore, whereas in a normally cooled compressor the temperature of the fluid entering the compressor and leaving the cooler is substantially constant for any load and is conveniently atmospheric temperature, in a compressor according to the invention, if atmospheric temperature were approached at full load, the cooler would be required to cool the fluid to sub-atmospheric temperatures at part load. In order to avoid such an undesirable expedient, therefore, the compressor may be arranged to operate on the basis that the temperature of the fluid at the cooler outlet is approximately atmospheric at the lowest required operational load and is progressively higher at progressively greater loads.

In consequence, such a compressor, operating at full load with inter-stage cooling to a temperature considerably above atmospheric, will be less eftective than a comparable compressor having normal inter-cooling; i. e. eifectiveness is sacrificed for stability.

However, calculations based on the conceptions outlined in the foregoing as they apply to two comparable gas turbine power plants, one having a compressor intercooled in accordance with the invention and the other having a normally cooled compressor, have indicated that the specific output of the former plant at full load is only 11 per cent. less than the corresponding output of the latter, this being accompanied by complete stability down to 36 per cent full load in the first case compared with 70 per cent in the second.

in order that the invention may be more clearly understood it will now be described by way of example in its application to an axial flow rotary air compressor comprising two stages of multiple-row blading mounted in tandem on a single shaft and constituting a component of a gas turbine power plant. Figures 2 to 9 of the accompanying drawings illustrate diagrammatically the comparative effects of varying degrees of inter-stage cooling on the characteristics of the machine, and Figures ii) and 11 illustrate practical embodiments for carrying out the method of the invention.

in the drawings:

Figures 2, 3 and 4 show graphically comparisons of the ratio of the densities, A, of compressed to uncompressed air (ordinates) at various axial distances, L (abscissa), through each of three compressors differing in respect of their inter-stage cooling; the origin in each case corresponds to the compressor intake where the density ratio is unity. 7

Figures 5, 6 and 7 show graphically comparisons of the axial velocity of the air, V (ordinates), at corresponding axial distances, L (abscissa), through each 0 3 the same three compressors.

Figures 8 and 9 show graphically comparisons of the respective characteristics of the low and high pressure stages of compressors and represent pressure ratio of the stage R (ordinates), plotted against mass flow, M (abscissa).

Figures 10 and 11 show diagrammatically two alternative arrangements of compressors embodying the invention.

With reference to the drawings, the effect of running an uncooled compressor on part load at a rotor speed below the designed speed is illustrated in Figure 2, in which the density ratio of the air throughout the compressor at the full load condition and at an arbitrary part load condition are respectively represented by the full line F and the chain dotted line P. Since at part load the pressure ratios of compressed to uncompressed air are progressively reduced, throughout the compressor, from their full load values, the air density ratios are similarly progressively reduced and in consequence, the axial velocity of the air through the compressor varies on reduction of load. A variation in axial velocity at part load is not wholly undesirable since, at the reduced rotor blade speed, the axial air velocity should be proportionately reduced in order to preserve the correct aerodynamic flow through the rotor blades; the actual variation which occurs, however, does not produce such a result and is, therefore, undesirable, and may cause surging. This effect is illustrated by Figure 5 in which the full line F represents the axial air velocity at the full load design condition and the intermittently broken line D represents the axial air velocity requi ed at an arbitrary part load to maintain, in conjunction with the reduced rotor blade speed, the correct aerodynamic flow, whilst the chaindotted line P represents the axial air velocity actually occurring at part load. As will be seen the velocity gradient through the compressor at part load is greater than is desirable.

In the case of a similar compressor having normal intercooling (down to a substantially constant temperature at all loads) the effects of such cooling on air density ratios and axial velocities are illustrated by Figures 3 and 6 respectively in which the conventions adopted to represent full and part load density ratio curves, and full load, and desired and actual part load velocity curves are respectively the same as for Figures 2 and 5. In this case, as

iii)

air flows through the cooler, its density ratio is sharply increased, but it will be observed that the change in density ratio caused by the cooler at part load is less than that at full load, due to the lower temperature reduction involved in normal inter-cooling at part loads, as previously described. Referring to Figure 6, it is assumed that the compressor is so designed that at full load the axial air velocity at the low pressure stage outlet is the same as at the high pressure stage inlet, and the full load velocity curve F is therefore a continuous line and similarly the velocity curve D required for correct aerodynamic how at the part load speed is also a continuous line. The effect of normal inter-cooling on the air at part load, however, due to the reduced change in air density ratio then induced, is to accelerate it between the compressor stages, with the result that the curve of actual axial air velocity P departs from that desired, D, to a greater extent than in the case of the uncooled compressor.

l1" now a compressor is considered having inter-cooling in accordance with the invention but otherwise similar to those previously considered, the curves shown in Figures 4 and 7 are obtained for conditions corresponding to those of the curves in Figures 2, 3 and 5, 6 respectively (the same conventions again applying for the various conditions). Referring to Figure 4 it will be seen that the effect of inter-cooling at part load is now to cause a greater increase in the air density ratio than at full load. In Figure 7 it is again assumed that the axial air velocities at the low pressure stage outlet and the high pressure stage inlet are arranged to be the same and that the full load and desired part load velocity curves are continuous lines. The effect of inter-cooling at part load in this case, however, due to the increase in density ratio being greater at part load than at full load, is to decelerate the air between the compressor stages, and thus the curve of axial air velocity P more nearly approaches the desired curve D than in either of the preceding cases.

The operating line characteristics (of the form illustrated generally in Figure l) corresponding to the three compressor cycles considered in the foregoing when constituting components of a gas turbine power plant, for low and high pressure stages respectively, re compared in Figures 8 and 9. in Figure 8, the curve U represents the operating line of the uncooled compressor in reistion to the surge line S, and the curve l\' that of the normally cooled compressor; the curve V1 represents the operating line of a compressor having var able intercooling in accordance with the invention. The curves V2 and V3 are intended to represent other operating lines obtainable by selecting rates of variation between intercooling and load other than that correspo iding to V1 when designing the machine. The line T represents the manner in which the operating lines V1, V2 or V3 would continue after reaching the lowest convenient temperature to whici inter-stagc cooling can be effected at the minimum operating load of the compressor, i. e. in the normal case, the atmospheric temperature line.

The high pressure stage characteristic is nearly identical for each cycle and is shown in Figure 3 as a sin lc curve, the point H being the designed compressor output condition.

The calculations previously referred to were carried out for a two-stage single shaft rotary axial compressor, and the final temperature of air leaving the cooler was taken as 350 K. at full load varying down to 293 at 30 per cent of full load.

In the practical application of the invention i is proposed that the effectiveness of the cooler should be varier by means of a control arranged to make the rate of low of coolant through the inter-cooler dependent upon the compressor load; the control should be responsive to changes in related variables, related to compressor load. such as, for instance, rotational speed, how, or the compressor delivery pressure to the cooler.

Various constructional methods of enecting such control present themselves; in one example, shown diagrammatically in Figure 10, a compressor having a low pressure stage 1 and a high pressure stage 2 mounted on the same shaft 3, is provided with an inter-stage cooler 4. A centrifugal governor 5 is driven through bevel gears 6 from the shaft 3 and is arranged to operate a link 7 which is connected to the valve 8 which controls the flow of coolant through the cooler 4. The valve 8 reduces or increases the flow of coolant upon respective upward or downward movement of link '7, that is, upon respective increase or decrease of the compressor speed and load. Any other type of speed governor could of course, be used in conjunction with suitable electrical, hydraulic, or mechanical servo mechanism to control the valve. In another form of control, shown diagrammatically in Figure 11, the compressor stages, interstage cooler and coolant control valve are as in Figure 10, but in this case the valve 8 is operated by a pressure responsive capsule 9 which is connected by a pipe 10 to the outlet of the low pressure stage 1 of the compressor. The valve 8 in Fig. 11 reduces or increases the flow'of coolant upon respective expansion or contraction of the pressure capsule 9, that is upon respective increase or decrease of the fluid pressure at the outlet of the low pressure compressor stage 1 and of the load on that stage. Although the embodiments of Figures 10 and 11 show the compressor stages mounted on the same shaft and include a valve for controlling the rate of supply of the coolant fluid, the application of the invention is not necessarily restricted solely to compressors having their stages mounted co-axially and for rotation at the same speed, or to metering the flow of a coolant fluid solely by a valve.

I claim:

1. A plant incorporating an elastic fluid rotary dynamic compressor having at least two stages arranged in series with respect to the flow of fluid therethrough, means for cooling said fluid arranged in intermediate series with said compressor stages, means affording a coolant supply to said cooling means, and means for controlling said coolant supply as a continuous function of the compress-3r load which coolant control means are so proportioned that the ratio of densities of the fluid after and before cooling respectively is increased when the compressor load decreases and is decreased when the compressor load increases.

2. A plant according to claim 1, wherein said means for controlling said coolant supply controls the rate of flow of said coolant supply and includes means responsive to the delivery pressure of said low pressure compressor stage to vary said rate of flow of said coolant supply.

3. A plant according to claim 1, wherein said means for controlling said coolant supply controls the rate of flow of said coolant supply and includes means responsive to rotational speed of the compressor stage delivering the fluid to be cooled.

References Cited in the file of this patent UNITED STATES PATENTS 1,580,435 Holdsworth Apr. 13, 1926 1,719,024 Replogle July 2, 1929 1,838,259 Hull Dec. 29, 1931 2,026,293 Wineman Dec. 31, 1935 2,074,803 Noble Mar. 23, 1937 2,078,956 Lysholm May 4, 1937 2,225,310 Lindhagen et al. Dec. 17, 1940 2,396,484 Allen et a1. Mar. 12, 1946 FOREIGN PATENTS 347,853 Germany Jan. 26, 1922 929,508 France Dec. 30, 1947 

